Power steering gear valve

ABSTRACT

A low noise rotary valve for a hydraulic power steering gear offering flexibility in the design of the boost characteristic has inlet ports to receive hydraulic fluid from a pump, return ports to return hydraulic fluid to the pump, and cylinder ports to communicate the hydraulic fluid to left and right-hand cylinder chambers of the hydraulic power steering gear. The valve has an input-shaft with a plurality of axially extending grooves separated by lands. A sleeve that has in its bore an array of axially extending slots that circumferentially align with the lands on the input-shaft is journalled on the input-shaft. A torsion bar residing in a bore of the input-shaft compliantly connects the input-shaft to a driven member. The interfaces between the co-acting input-shaft grooves and sleeve slots define axially extending orifices which open and close when relative rotation occurs between the input shaft and the sleeve. These orifices are ported as a network such that they form primary and secondary hydraulic Wheatstone bridges each having right and left-hand inlet orifices and return orifices. Hydraulic flow from the primary bridge is hydraulically communicated to the return port via a primary return path and the hydraulic flow from the secondary return path is hydraulically communicated to the return port via a secondary return path. The secondary return path contains an annular restriction having a cross-section to flow which has a high aspect ratio. Thus, the return path of each secondary bridge provides a restriction in hydraulic flow to the upstream orifices that causes a back pressure to be applied to the secondary upstream orifices. This back pressure significantly suppresses the generation of cavitation noise in the upstream secondary orifices.

TECHNICAL FIELD

This invention relates to rotary valves such as are used in hydraulicpower steering gears for vehicles. More particularly the presentinvention provides low noise rotary valves offering flexibility in thedesign of the boost characteristic.

BACKGROUND ART

Such rotary valves typically include an input-shaft which incorporatesin its outer periphery a plurality of blind-ended, axially extendinggrooves separated by lands. Journalled on the input-shaft is a sleevehaving in its bore an array of axially extending blind-ended slotscircumferentially aligned with the lands on the input-shaft. Theinterfaces between the coacting input-shaft grooves and sleeve slotsdefine axially extending orifices which open and close when relativerotation occurs between the input-shaft and the sleeve. The sides of theinput-shaft grooves are contoured so as to provide a specific orificeconfiguration and are referred to as metering edge contours. Theseorifices are ported as a network such that they form sets of hydraulicWheatstone bridges which act in parallel. Such hydraulic Wheatstonebridges are analogous in operation to conventional electrical Wheatstonebridges.

Drilled passages in the input-shaft and sleeve, together withcircumferential grooves in the periphery of the sleeve, serve tocommunicate oil between the grooves in the input-shaft and the slots inthe sleeve, a hydraulic pump, and right-hand and left-hand hydraulicassist cylinder chambers incorporated in the steering gear.

A torsion bar incorporated in the input-shaft serves to urge theinput-shaft and sleeve towards a neutral, centred position when no powerassistance is required. When input torque is applied by the driver tothe steering wheel, the torsion bar deflects, causing relative rotationof the sleeve and input-shaft from the neutral position. This so called"valve operating angle" imbalances the sets of hydraulic Wheatstonebridges and hence causes a differential pressure to be developed betweenthe right-hand and left-hand cylinder chambers. The "boostcharacteristic" of the rotary valve, that is the functional relationshipbetween the above mentioned input torque and differential pressure, islargely determined for a given steering gear application by the geometryof the metering edge contours.

Traditionally the network of orifices in a rotary valve employ 2, 3 or 4Wheatstone bridges, necessitating respectively 4, 6 or 8 input-shaftgrooves and sleeve slots. Each Wheatstone bridge comprises a right-handand a left-hand oil flow path, henceforth termed "limbs", and eachright-hand and left-hand limb in turn comprises upper and lowerportions. The upper and lower portions of each right-hand and left-handlimb meet respectively at a point of connection to the right-hand andleft-hand cylinder chamber, henceforth termed the right-hand andleft-hand "cylinder ports" of the valve.

In the neutral position of the rotary valve, oil from the hydraulic pumpdivides and enters each Wheatstone bridge at the valve "inlet port". Atthis point flow further divides and enters the upper right-hand andleft-hand limbs, each containing an "inlet orifice". After being meteredthrough such inlet orifices, oil is communicated to the respectivecylinder ports and to the respective interconnection to the lower limbs.Depending on the intercylinder flow rate drawn by the motion of thepiston in the cylinder chamber, oil continues to flow through the lowerright-hand and left-hand limbs, metering through a "return orifice" ineach limb, and recombining immediately upstream of the "return port" ofthe rotary valve.

The network of two inlet orifices and two return orifices, constitutingeach Wheatstone bridge, is ported in the rotary valve such that, for agiven relative angular displacement of the input-shaft and sleeve fromtheir neutral position, mutually opposite orifices on each Wheatstonebridge simultaneously close or open. For example, the left-hand inletand right-hand return orifices both close (ie. increase in restrictionto oil flow) while the right-hand inlet and left-hand return orificesboth open (ie. decrease in restriction to oil flow). According toclassical Wheatstone bridge theory, for a given oil flow through eachWheatstone bridge, a differential pressure is therefore developedbetween the right-hand and left-hand cylinder ports, providing thenecessary level of power assistance for each value of valve operatingangle.

The general method of operation of such traditional rotary valves iswell known in the prior art of power steering design and described ingreater detail in U.S. Pat. No. 3,022,772 (Zeigler et al), commonly heldas being the "original" patent describing the rotary valve concept.Rotary valves of this format will henceforth be termed "direct modevalves" since all Wheatstone bridges within the valve incorporate directhydraulic communication to the cylinder ports.

Rotary valves are nowadays regularly incorporated in firewall-mountedrack and pinion steering gears, and in this situation, any noises suchas hiss emanating from the valve are very apparent to the driver. Hissresults from cavitation of the hydraulic oil as it flows in the orificesdefined by the metering edge contours and the adjacent edges of thesleeve slots, particularly during times of high pressure operation ofthe valve such as during parking, where differential pressures of 8-10MPa or more can be generated. It is well known in the art of powersteering valves that an orifice is less prone to cavitation if theassociated metering edge contour has a high aspect ratio of axial lengthto radial depth, thereby constraining the oil to flow as a thin sheet ofconstant depth along the full axial extent of one metering edge contourand if, furthermore, the flow of oil is evenly divided amongst severalmetering edge contours ported to act in parallel, so further effectivelyreducing the flow of oil that may flow through any one orifice. It isalso well known that cavitation is less likely to occur if the meteringedge contour, where it intersects the outside diameter of theinput-shaft, is nearly tangential thereto, hence constituting a shallowchamfer typically inclined at an angle of between 4 deg and 8 deg.

Such shallow chamfers have been widely used in rotary valves for noisesuppression over the last 20 years. In order to achieve the necessarydepth and form accuracy, these chamfers are normally ground in specialindexable or cam-type grinding machines resulting in long overall cycletimes, relatively expensive capital equipment, and hence high overallmanufacturing cost.

Another requirement which is increasingly becoming accepted for thedesign of rotary valves is the need for a linear boost characteristic.During vehicle cornering, it is advantageous that a substantially linearrelationship exists between the driver's input torque and thedifferential pressure associated with such a cornering manoeuvre. Thisleads to the sensation of "progression" in the power assistance andmaximises steering feel in such critical situations. Associated with therequirement for a linear boost characteristic, it is also highlydesirable to maximise the extent of the linear region before the maximumparking torque (and hence parking pressure) is reached. Thisnecessitates a fast transition or "turn-around" of the linear boostcharacteristic to a region of much steeper slope associated with thehigher differential pressures used for parking. For a given slope andextent of the linear boost characteristic of the rotary valve duringcornering, the torque required to be exerted by the driver duringparking is therefore minimised.

Chamfer type metering edge contours can, to some limited degree,generate a low noise linear boost characteristic if the chamfer isdesigned as a scroll, as disclosed in U.S. Pat. No. 5,267,588 (Bishop etal), or as a series of flat facets, as disclosed in U.S. Pat. No.4,460,016 (Haga et al). However, in both these cases, the extent of thelinear boost region is relatively short and the transition to thesteeper parking region of the boost characteristic is prolonged, andtherefore not optimal in terms of minimising parking torque.

Another technique, well known in the art, for suppressing valve noise inpower steering valves is the application of back pressure to anotherwise cavitating orifice, thereby raising pressures within theorifice above the vapour pressure of the hydraulic oil and hencepreventing the onset of cavitation. Chamfer style metering edge contoursneed not necessarily be used on the input-shaft if this alternativemethod of noise suppression is employed. Much steeper and axiallyshorter metering edge contours can in fact be used, contours which wouldotherwise be excessively noisy in the absence of such back pressure.Such steeper and generally more complex shaped metering edge contourscan be manufactured by coining, roll-imprinting or traditional hobbingmethods and, if appropriately designed, can generate the previouslydescribed desirable linear boost characteristic with a fast turn-around.

U.S. Pat. No. 4,335,749 (Walter) shows a direct mode valve incorporatingan extra (second) orifice in the lower portions of the left-hand andright-hand limbs of each bridge. This orifice progressively closes withincreasing valve operating angle until a constant orifice area isreached which, based on the flow through the limb, applies apredetermined back pressure to the closing upstream inlet orifice. Sucha valve format is based on 6 orifices per bridge and, if 3 bridges areemployed in the valve, requires 9 input-shaft grooves and 9 sleeveslots. If 4 bridges are employed (as in the case of a traditional 8groove/slot rotary valve), 12 input-shaft grooves and 12 sleeve slotsare required. This format is therefore non-standard and requires extramanufacturing cost.

Further however, experiments have shown that elimination of cavitationnoise in an orifice generating 10 MPa (say) differential pressurerequires a downstream back pressure to be applied which is as much as 1MPa or more . If such high levels of back pressure were generated by thereturn orifices according to the methodology disclosed in U.S. Pat. No.4,335,749 (Walter), this back pressure would raise the inlet pressurerequired to be supplied by the hydraulic pump by the same 1 MPa, withoutany of this additional pressure being applied differentially at thecylinder chamber. This is because this direct mode valve arrangementcontains cylinder port connections in every bridge and the returnorifice used for back pressure generation is downstream of suchconnections. The 1 MPa increase in valve inlet pressure would be totallywasted in terms of generating power assistance force and would simplyraise the operating pressure of the hydraulic pump. The latter situationis highly undesirable since energy loss in the hydraulic system isproportionally increased. Also pump noise, leakage and potentialhydraulic line failure all become bigger problems as the pump reliefvalve setting is necessarily increased to accommodate the increasedvalve operating pressure, for example from 10 MPa to 11 MPa in thiscase.

For this reason the practical level of back pressure that can be appliedby the return orifice according to the above prior art invention islimited to about 300-400 kPa, well short of the 1 MPa or more needed tosubstantially eliminate cavitation noise through the operating pressurerange of the power steering valve.

Another class of rotary valve, henceforth termed "bypass mode valves",is quite distinct from the class of direct mode valves earlierdescribed. Bypass mode valves also utilise parallel arrangements ofWheatstone bridges, however not all bridges in this case contain ahydraulic connection to a cylinder port between the inlet and returnorifices. The bridges which employ a cylinder connection will henceforthbe termed "primary bridges" and those which don't employ a cylinderconnection termed "secondary bridges". In the latter case the left-handand right-hand limbs contain one or more inlet and return orifices butwith no interposed cylinder port connection. In this manner, for certainvalve operating angles, hydraulic oil at least partially by-passes theprimary bridge(s) which incorporate the connection to the cylinder.

Such bypass mode valves were first put forward for speed sensitive powersteering applications. For example, arrangements described in U.S. Pat.Nos. 4,570,735 (Duffy) and 4,570,736 (Waldorf) and Japanese Patent04-031175 (Suzuki et al) involve a bypass mode valve with anelectronically modulated variable orifice residing in the inlet to thesecondary bridges and modulated as a function of vehicle speed. Otherlater arrangements such as shown in Japanese Patent 02-306878 (Suzuki)and U.S. Pat. No. 5,092,418 (Suzuki et al) use an electronicallymodulated variable orifice residing in the return line from thesecondary bridges. In such speed sensitive applications, the degree ofbypass of hydraulic oil through the secondary bridges is used to controlthe boost characteristic as a function of vehicle speed.

Bypass mode valves have also been utilised in a non-speed sensitiveformat to improve the linearity and produce a fast turn-around of theboost characteristic for valves employing chamfered metering edgecontours. For example Japanese Patents 04-031176, 05-042880 and06-278623 (all Suzuki et al) and U.S. Pat. No. 4,470,432 (Kervagoret)show orifice networks very similar to the abovementioned speed sensitiveapplications except that the electronically modulated variable orificeis now a fixed "drill-hole" style orifice either upstream or downstreamof the secondary bridges. In situations where relatively conventionalmetering edges are used in the orifices of the secondary bridges, sucharrangements will tend to be noisy for two reasons. Firstly the very lowaspect ratio of the fixed orifices (ie. unity for a drill hole) will bea source of cavitation for the relatively high oil flows involved.Secondly in these arrangements, for high valve operating pressures, allpump flow is communicated to the return port via only two stages ofpressure drop: the restrictions offered by relevant closing secondaryorifice and the fixed orifice (or vice versa).

U.S. Pat. No. 4,577,660 (Haga) shows an 8 slot by-pass valve againintended to produce a linear boost characteristic with a fast turnaround. In this case the secondary inlet orifices are overlapped and infact closed on-centre, their sudden opening off-centre intended toproduce the required discontinuity in the boost characteristic. Howeversuch an arrangement, with a substantial portion of the valve closedon-centre, would naturally exhibit higher than usual back pressureon-centre and would therefore be energy inefficient.

Japanese Patent 04-292265 (Suzuki et al) shows a relatively complexbypass mode valve employing 10 input-shaft grooves and corresponding 10sleeve slots. An extra orifice is positioned in the lower portion ofeach secondary bridge limb and, as it closes, provides a similar flowredistribution function to the earlier mentioned fixed orifice. Such avalve arrangement is expensive because of the larger quantity ofinput-shaft grooves and sleeve slots, and the associated interconnectingporting (eg. drill holes) to hydraulically communicate suchslots/grooves. Moreover 10 input-shaft grooves and 10 sleeve slots aredifficult to package using a standard input-shaft outside diameter (orcorresponding sleeve inside diameter), typically in the range 19.0-22.5mm, and yet still retain sufficient inter and intra slot/groove spacingto accommodate such interconnecting porting.

However the nature of the diversion of oil flow in bypass mode valvesbetween the primary and secondary bridges means that, when such valvesgenerate a large differential pressure at the cylinder, essentially onlythe secondary bridges transmit any oil flow. This means that theindividual orifices in the secondary bridges tend to be prone tocavitation noise even if shallow chamfers are employed as the meteringedge contours according to the prior art. Japanese Patent 05-310136(Suzuki et al) proposes to reduce this problem by employing anelectronically modulated variable orifice positioned at the return portof a bypass mode valve, this variable orifice controlled to produce arestriction (and hence generate back pressure) as a function of thesensed inlet pressure to the valve. For reasons earlier described, suchan arrangement is energy inefficient and, moreover in this case, addssignificant cost to the power steering system.

Nevertheless, bypass mode valves do offer a major advantage over directmode valves in that, any back pressure applied within the secondarybridge network to suppress cavitation noise generated by the respectiveorifices does not raise the overall inlet pressure to the valve, assupplied by the hydraulic pump, for a given differential pressureapplied at the cylinder. Hence such back pressure is not wasteful interms of energy and in fact is usefully used to generate some portion ofthe power assistance at the cylinder. There is therefore no need to usehigher pump relief valve settings and the previously referred to largelevels of back pressure (eg. 1 MPa) can be theoretically utilised tosubstantially eliminate cavitation noise without any major disadvantagein terms of valve function.

The first and second aspects of the present invention are directed atutilising the above mentioned benefits of bypass mode valves, and yetprovide low levels of cavitation noise in a rotary valve withoutnecessarily increasing the number of input-shaft grooves or sleeveslots. Another aim is to permit such low levels of cavitation noise tobe achieved using metering edge contours of the more steeply slopingvariety earlier referred to. Such metering edge contours can be producedby coining, roll-imprinting or hobbing and not only offer significantcost savings compared to shallow chamfers which generally must beground, but also enable much more flexibility in the design of the boostcharacteristic, particularly the provision of a linear boostcharacteristic with a fast turn-around. Also such steeply slopingmetering edge contours can generally be designed to be axially shorterthan the comparable shallow chamfer, enabling the overall rotary valvepackage to be likewise shortened.

The first aspect of the present invention consists in a rotary valve fora hydraulic power steering gear comprising a valve housing having aninlet port to receive hydraulic fluid from a pump, a return port toreturn hydraulic fluid to the pump, and cylinder ports to communicatehydraulic fluid to left and right-hand cylinder chambers of the powersteering gear, the valve also comprising an input-shaft having in itsouter periphery a plurality of axially extending grooves separated bylands, a sleeve journalled on said input-shaft, said sleeve having inits bore an array of axially extending slots circumferentially alignedwith the lands on the input-shaft, the interfaces between the coactinginput-shaft grooves and sleeve slots defining axially extending orificescontrolling fluid flow within the valve, the orifices opening andclosing when relative rotation occurs between the input-shaft and sleevefrom a neutral position, the orifices being ported as a network suchthat they form one or more primary and one or more secondary hydraulicWheatstone bridges arranged in parallel, each said bridge comprising twolimbs hydraulically communicating the inlet and return ports, each saidlimb containing an inlet orifice hydraulically communicating to theinlet port and a return orifice hydraulically communicating to thereturn port, the magnitude of the hydraulic flow through each bridgevarying in accordance with the restriction offered by the respectiveinlet and return orifices in that bridge, the limbs of the primarybridge incorporating means providing hydraulic communication to one ofthe cylinder ports at a point of interconnection of the respective inletand return orifices in that limb, the limbs of the secondary bridge notincorporating means providing hydraulic communication to the cylinderports, characterised in that the return orifice in each limb of saidsecondary bridge is formed by a metering edge contour on the edge of thesecondary return groove associated with said return orifice, saidmetering edge contour circumferentially overlapping the adjacent sleevebore land when the rotary valve is in its neutral position to such anextent that said return orifice provides a restriction to hydraulic flowas the upstream inlet orifice in the same limb closes for all valveoperating angles from said neutral position, said return orificeapplying a back pressure to said upstream inlet orifice, said backpressure being sufficient to significantly suppress the generation ofcavitation noise in said inlet orifice.

It is preferred that a substantially constant restriction area isprovided by said return orifice as the upstream inlet orifice in thesame limb closes for all valve operating angles from said neutralposition.

It is preferred that the input-shaft metering edge contour employed insaid return orifice is formed in cross-section such that a region oflocally reduced metering edge depth lies in the overlapped region of thecoacting input-shaft metering edge contour and adjacent sleeve boreland, that is in the region lying radially inside the adjacent sleeveland. Said substantially constant restriction area provided by saidreturn orifice can be considered as constituting a hydraulic throatwhich serves to significantly suppress cavitation noise or turbulence asthe hydraulic oil flows past the adjacent sleeve edge and enters thisreturn orifice.

It is preferred that cavitation and other flow noise can be furtherreduced by raising the back pressure downstream of the secondary returnorifice.

It is preferred that hydraulic flow from the primary bridge ishydraulically communicated to the return port via a primary return pathand the hydraulic flow from the secondary bridge-is hydraulicallycommunicated to the return port via a secondary return path, arestriction existing in the secondary return path.

In first and second embodiments it is preferred that the secondaryreturn path passes through the bore of the input-shaft.

In a first embodiment it is preferred that the radial holes whichhydraulically communicate the secondary return groove to the input-shaftbore are reduced in diameter, thereby generating back pressure in thesecondary return groove downstream of the secondary return orifice.

The capability of these radial holes to apply such back pressure withoutthemselves causing a noise problem can be further enhanced if conical ortapered entries are employed on these holes. This can be readily andcheaply achieved via a step form on the drill used to machine theseholes or by laser erosion.

In the case of this first embodiment, it is also preferred that theprimary return path also passes through the bore of the input-shaft.However the radial holes which hydraulically communicate the primaryreturn grooves to the input-shaft bore are sufficiently large indiameter in this case that no substantial restriction is generated.

In a second embodiment, and also in a later referred to fifthembodiment, the restriction existing in the secondary return path ispreferably annular in geometry. It is preferred that this annularrestriction has a cross-section to flow which has a high aspect ratio,in order to suppress its generation of cavitation noise. In this secondembodiment it is preferred that only the secondary return path passesthrough the bore of the input-shaft. Oil entering the input-shaft borevia the aforementioned radial holes is restricted using a diametricallyenlarged portion on the torsion bar. This enlarged portion is arrangedto have a small radial clearance with respect to the input-shaft bore,hence creating an annular restriction for hydraulic oil as it flowsaxially in this bore towards the return port of the valve housing.

According to this embodiment, hydraulic oil from the primary bridge isported directly to the return port so that it is not required to flowthrough the input-shaft bore, and hence is not subject to thisadditional restriction. This is achieved by axially extending theinput-shaft grooves associated with the primary return orifices in theform of channels, allowing hydraulic oil flow in the primary bridge toexit directly axially from these grooves through these channels.

The diametrically enlarged portion of the torsion bar can be integrallymachined as part of the torsion bar during its manufacture. However, inorder to maximise the working length of the reduced diameter portion ofthe torsion bar, and hence lower the maximum stress endured by thetorsion bar for a given working diameter and torsional spring rate, thediametrically enlarged portion on the torsion bar is preferably formedas an annular bush which is plastic moulded around the metallic portionof the torsion bar as a separate subsequent operation. The plasticmaterial must be chemically resistant to hydraulic oil and is preferablyan engineering plastic such as Delrin® or Lurathane®.

If the annular bush is made to additionally extend axially such that itoverlaps the secondary return radial holes in the input-shaft, the useof such compliant plastic material for this bush has been found toassist dampening the hydraulic turbulence noise generated by thesecondary return oil as it flows radially into the input-shaft bore andthence necessarily turns perpendicularly to continue flowing axiallydown this bore.

In a third embodiment it is preferred that the secondary return pathdoes not pass through the bore of the input-shaft. Axially extendingsecondary return channels are formed in the sleeve bore which arecircumferentially aligned with the secondary return grooves. Thechannels extend to the axial extremity of the sleeve bore and arrangedto communicate hydraulic fluid to the return port. The radial depth ofthe channels is small, thereby interacting with the adjacent outsidediameter of the input-shaft to form a high aspect ratio restriction inthe secondary return path downstream of the secondary return grooves. Itis preferred that at least one secondary return channel extends to bothaxial extremities of the sleeve bore.

It is also preferred that axially extending primary return channels areformed in the sleeve bore and arranged to be circumferentially alignedwith the primary return grooves. These additional channels also extendto the axial extremity of the sleeve bore and are arranged tocommunicate hydraulic fluid to the return port. It is preferred that theradial depth of the primary return channels is larger than that of theaforementioned shallow secondary return channels since no restriction isrequired to be generated in the primary return path. It is alsopreferred that at least one of the primary return channels extends toboth axial extremities of the sleeve bore.

For reasons of ease of manufacture, it is preferred that all primary andsecondary return channels extend to both axial extremities of the sleevebore, enabling all such channels to be formed with a single multi-toothbroaching tool.

In a fourth embodiment it is also preferred that the secondary returnpath does not pass through the bore of the input-shaft. The secondaryreturn grooves are axially extended as shallow, high aspect ratiochannels formed via their interaction with the adjacent sleeve bore.These channels extend to the axial extremity of the sleeve bore, therebyproviding a restriction in the secondary return path. It is alsopreferred that the primary return grooves are similarly axially extendedas radially deeper channels to facilitate a relatively unrestrictedprimary return path.

In a fifth embodiment is is also preferred that the secondary returnpath does not pass through the bore of the input-shaft. The secondaryreturn grooves are axially extended in at least one direction tocommunicate with an annular cavity formed by the interaction of areduced diameter portion of the input-shaft outer periphery and thesleeve bore. The annular cavity acts as a manifold to gather secondaryreturn oil flow, which is then communicated via an annular restrictionto the return port. The annular restriction is preferably generated by apredetermined small radial clearance existing between the abovementioned reduced diameter portion of the input-shaft and the insidediameter of a radially inwardly extending portion of the sleeve bore.Preferably the radially inwardly extending portion of the sleeve bore isformed as an accurately internally and externally sized annularpressed-metal cup which is press-fitted inside the sleeve skirt to sealagainst the axial extremity of the sleeve bore. Preferably thepredetermined radial clearance is such that the resulting annularrestriction has a high aspect ratio in order to suppress its generationof cavitation noise. It is also preferred that the primary return pathpasses through the bore of the input-shaft in a similar manner to thatdescribed in reference to the first embodiment, thereby bypassing theannular restriction en-route to the return port.

It is preferred that the rotary valve has eight input-shaft grooves.

It is preferred that the rotary valve has eight sleeve slots.

The second aspect of the present invention consists in a rotary valvefor a hydraulic power steering gear comprising a valve housing having aninlet port to receive hydraulic fluid from a pump, a return port toreturn hydraulic fluid to the pump, and cylinder ports to communicatehydraulic fluid to left and right-hand cylinder chambers of the powersteering gear, the valve also comprising an input-shaft having in itsouter periphery a plurality of axially extending grooves separated bylands, a sleeve journalled on said input-shaft and rotationally securedto a driven member, said sleeve having in its bore an array of axiallyextending slots circumferentially aligned with the lands on theinput-shaft, the interfaces between the coacting input-shaft grooves andsleeve slots defining axially extending orifices controlling fluid flowwithin the valve, the orifices opening and closing when relativerotation occurs between the input-shaft and sleeve from a neutralposition, a torsion bar residing in a bore of the input-shaftcompliantly connecting the input-shaft and driven member, and arrangedto urge the sleeve and input-shaft to the neutral position, the orificesbeing ported as a network such that they form one or more primary andone or more secondary hydraulic Wheatstone bridges arranged in parallel,each said bridge comprising two limbs hydraulically communicating theinlet and return ports, each said limb containing an inlet orificehydraulically communicating to the inlet port and a return orificehydraulically communicating to the return port, the magnitude of thehydraulic flow through each bridge varying in accordance with therestriction offered by the respective inlet and return orifices in thatbridge, the limbs of the primary bridge incorporating means providinghydraulic communication to one of the cylinder ports at a point ofinterconnection of the respective inlet and return orifices in thatlimb, the limbs of the secondary bridge not incorporating meansproviding hydraulic communication to the cylinder ports, characterisedin that hydraulic flow from said primary bridge is hydraulicallycommunicated to the return port via a primary return path and thehydraulic flow from the secondary bridge is hydraulically communicatedto the return port via a secondary return path, an annular restrictionexisting in the secondary return path.

It is preferred that the annular restriction existing in the secondaryreturn path has a cross-section to flow which has a high aspect ratio.

It is preferred that the aspect ratio be greater than 10.

It is preferred that one but not both of the primary or secondary returnpaths passes through the bore of the input-shaft.

In a first embodiment it is preferred that the secondary return pathpasses through the bore of the input-shaft and the annular restrictionis formed within this bore.

It is preferred that the annular restriction formed in the bore of theinput-shaft is generated by virtue of a small radial clearance existingbetween a diametrically enlarged portion of the torsion bar and theinput-shaft bore.

It is preferred that hydraulic flow from the primary bridge is directlycommunicated to the return port via channels formed as an axialextension of the input-shaft grooves associated with the primary returnorifices. Because this hydraulic flow is not communicated through theinput-shaft bore, it is not subject to the abovementioned annularrestriction.

Various preferred embodiments are possible for the geometry andconstruction of the diametrically enlarged portion of the torsion barand have already been described in reference to the first aspect of thepresent invention.

In a second embodiment is is preferred that the secondary return pathdoes not pass through the bore of the input-shaft and the annularrestriction is formed at the input-shaft/sleeve interface. The secondaryreturn grooves are axially extended in at least one direction tocommunicate with an annular cavity formed by the interaction of areduced-diameter portion of the input-shaft outer periphery and thesleeve bore. The annular cavity acts as a manifold to gather secondaryreturn oil flow, which is then communicated via an annular restrictionto the return port. The annular restriction is preferably generated by apredetermined small radial clearance existing between the abovementioned reduced diameter portion of the input-shaft and the insidediameter of a radially inwardly extending portion of the sleeve bore.Preferably the radially inwardly extending portion of the sleeve bore isformed as an accurately internally and externally sized annular pressedmetal cup which press-fitted inside the sleeve skirt to seal against theaxial extremity of the sleeve bore. Preferably the predetermined radialclearance is such that the resulting annular restriction has a highaspect ratio in order to suppress its generation of cavitation noise. Itis also preferred that the primary return path passes through the boreof the input-shaft in a similar manner to that described in reference tothe first embodiment of the first aspect of the present invention,thereby bypassing the annular restriction en-route to the return port.

In the case of both first and second embodiments of the second aspect ofthe present invention, hydraulic flow from the secondary bridge passesaxially through the relevant annular restrictions and hence applies aback pressure to all secondary orifices upstream of this restriction.The restriction area is therefore substantially constant and arranged toprovide sufficient back pressure to suppress the generation ofcavitation noise in these secondary orifices for all valve operatingangles.

It is preferred that the rotary valve has eight input-shaft grooves.

It is preferred that the rotary valve has eight sleeve slots.

BRIEF DESCRIPTION OF THE DRAWINGS

In order that the first and second aspects of the present invention maybe better understood, various embodiments thereof are now described, byway of example, with reference to the accompanying drawings, in which:

FIG. 1 is an axial cross-sectional view on plane I--I in FIG. 2 of arotary valve installed in a valve housing of a power steering gearaccording to a first embodiment of the first aspect of the presentinvention;

FIG. 2 is a cross-sectional view of the input-shaft and surroundingsleeve components of the rotary valve on plane II--II in FIG. 1;

FIG. 3 is an enlarged version of the upper half of the cross-sectionalview shown in FIG. 2, indicating the orifices in the primary andsecondary bridges;

FIG. 4 is the hydraulic "flow diagram" for the network of orifices shownin FIG. 3 corresponding to one primary and one secondary bridge inparallel ie. one half of the overall rotary valve hydraulic circuit;

FIG. 5 shows details of the primary inlet orifices;

FIG. 6 shows details of the primary return orifices;

FIG. 7 shows details of the secondary inlet orifices;

FIG. 8 shows details of the secondary return orifices;

FIG. 9 is a graph plotting the angular boost characteristic of therotary valve;

FIG. 10 is a graph plotting the flow division between the primary andsecondary bridges in the rotary valve as a function of valve operatingangle;

FIG. 11 is a graph plotting the flow division between the primary andsecondary bridges in the rotary valve as a function of differentialpressure;

FIG. 12 is a simplified version of the hydraulic "flow diagram" shown inFIG. 4, to assist in the understanding of valve operation in thecornering and parking regions of the boost characteristic;

FIG. 13 is a graph plotting back pressures Pb and Pr, developed byorifices 34b and 46 respectively in the rotary valve, as a function ofvalve operating angle;

FIG. 14 is a graph plotting back pressures Pb and Pr, developed byorifices 34b and 46 respectively in the rotary valve, as a function ofdifferential pressure;

FIGS. 15. a-d are detailed scrap views of region G in FIG. 3 showingvarious possible embodiments of the entry to radial hole 25;

FIG. 16 is an axial cross-sectional view on plane XVI--XVI in FIG. 18 ofa rotary valve installed in a valve housing of a power steering gear,according to a second embodiment of the first aspect of the presentinvention;

FIG. 17 is an axial cross-sectional view on plane XVII--XVII in FIG. 18of a rotary valve installed in a valve housing of a power steering gear,according to a second embodiment of the first aspect of the presentinvention;

FIG. 18 is a cross-sectional view on plane XVIII--XVIII in FIGS. 16 and17 of the input-shaft and surrounding sleeve components of the rotaryvalve, according to a second embodiment of the first aspect of thepresent invention;

FIGS. 19 a-d are enlarged scrap views of region E in FIGS. 16 and 26,showing various embodiments for the diametrically enlarged portion ofthe torsion bar;

FIG. 20 is a sectional view on plane XX--XX in FIG. 19c;

FIG. 21 is an alternative version of the first embodiment of the firstaspect of the present invention shown in FIG. 1, where the torsion barincorporates a surrounding plastic moulding;

FIG. 22 is an axial cross-sectional view on plane XXII--XXII in FIG. 24of a rotary valve installed in a valve housing of a power steering gear,according to a third embodiment of the first aspect of the presentinvention;

FIG. 23 is an axial cross-sectional view on plane XXIII--XXIII in FIG.24 of a rotary valve installed in a valve housing of a power steeringgear, according to a third embodiment of the first aspect of the presentinvention;

FIG. 24 is a cross-sectional view on plane XXIV--XXIV in FIGS. 22 and 23of the input-shaft and surrounding sleeve components of the rotaryvalve, according to a third embodiment of the first aspect of thepresent invention;

FIG. 25 shows details of the secondary return orifices according tofirst and second embodiments of the second aspect of the presentinvention;

FIG. 26 is an axial cross-sectional view on plane XXVI--XXVI in FIG. 28of a rotary valve installed in a valve housing of a power steering gear,according to a first embodiment of the second aspect of the presentinvention;

FIG. 27 is an axial cross-sectional view on plane XXVII--XXVII in FIG.28 of a rotary valve installed in a valve housing of a power steeringgear, according to a first embodiment of the second aspect of thepresent invention;

FIG. 28 is a cross-sectional view on plane XXVIII--XXVIII in FIGS. 26and 27 of the input-shaft and surrounding sleeve components of therotary valve, according to a first embodiment of the second aspect ofthe present invention;

FIG. 29 is an axial cross-sectional view of a rotary valve installed ina valve housing of a power steering gear, according to a fifthembodiment of a first aspect of the present invention, showing thesecondary return path;

FIG. 30 is an enlarged scrap view of a portion of FIG. 29 showingdetails of the secondary return path;

FIG. 31 is an axial cross-sectional view of a rotary valve installed ina valve housing of a power steering gear, according to a fifthembodiment of a first aspect of the present invention, showing theprimary return path;

FIG. 32 is an enlarged scrap view of a portion of FIG. 31 showingdetails of the primary return path;

FIG. 33 is an axial cross-sectional view of a rotary valve installed ina valve housing of a power steering gear, according to a secondembodiment of a second aspect of the present invention, showing thesecondary return path;

FIG. 34 is an enlarged scrap view of a portion of FIG. 33 showingdetails of the secondary return path;

FIG. 35 is an axial cross-sectional view of a rotary valve installed ina valve housing of a power steering gear, according to a secondembodiment of a second aspect of the present invention, showing theprimary return path; and

FIG. 36 is an enlarged scrap view of a portion of FIG. 35 showingdetails of the primary return path.

BEST MODES

FIGS. 1-15 and FIG. 21 refer to a first embodiment of the first aspectof the present invention. Referring to FIG. 1 valve housing 1 isprovided with pump inlet and return ports 2 and 3 respectively and rightand left-hand cylinder ports 4 and 5. Steering gear housing 6, to whichvalve housing 1 is attached, contains the mechanical steering elements,for example, a driven member in the form of pinion 7, journalled byneedle roller bearing 8 and provided with seal 9. The three main rotaryvalve elements comprise input-shaft 10, sleeve 11 journalled thereon,and torsion bar 12. Torsion bar 12 is secured by pin 13 to input-shaft10 at one end, and secured by swageing 14 to pinion 7 at the other.Torsion bar 12 also provides a journal for input-shaft 10 at overlappingportion 15. Sleeve 11 has an annular extension having therein hole 16engaging pin 17 extending radially from pinion 7.

Referring now also to FIG. 2, input-shaft 10 incorporates on its outsidediameter 20 eight axially extending, blind-ended grooves 18a-c separatedby lands 81: four grooves of the type indicated as 18a, two of the typeindicated as 18b, and two of the type indicated as 18c. Sleeve 11incorporates in its bore 21 an array of eight axially extending,blind-ended slots 19a-b separated by lands 82: four slots of the typeindicated as 19a and four of the type indicated as 19b. Slots 19a-b arecircumferentially aligned with lands 81 on input-shaft 10. Similarlygrooves 18a-c are circumferentially aligned with lands 82 on bore 21 ofsleeve 11. Metering edge contours are formed on the sides of all eightgrooves 18a-c and coact with the respective adjacent edges of slots19a-b to define sixteen axially extending orifices which open and closewhen relative rotation occurs between input-shaft 10 and sleeve 11.

Sleeve 11 is also provided on its outside periphery with three axiallyspaced circumferential grooves 22a-c separated by high pressure seals 23(see FIG. 1). Radial holes 24 and 25 in input-shaft 10 hydraulicallycommunicate grooves 18b and 18c respectively to bore 26 of in put-shaft10, whence return oil can flow back to the pump reservoir (not shown)via return port 3.

Radial holes 27 in sleeve 11 hydraulically communicate the remainingfour alternate grooves 18a of input-shaft 10 to the centralcircumferential groove 22b, and so to the supply from the hydraulic pump(not shown) via inlet port 2.

Radial holes 28 in sleeve 11 hydraulically communicate pairs of adjacentslots 19a of sleeve 11 to circumferential grooves 22a and 22c and thenceto the right-hand and left-hand cylinder chambers (not shown) viaright-hand cylinder port 4 and left-hand cylinder port 5 respectively.

The aforementioned sixteen axially extending orifices in the rotaryvalve are ported as a network such that they form a set of fourhydraulic Wheatstone bridges: two primary bridges residing in sectors 29of the valve and two secondary bridges residing in sectors 30. Theparallel action of the diametrically opposed bridges of the same typeensures that substantially zero net side force is produced on theinput-shaft due to the pressure distribution in the valve, minimisingfriction at the input-shaft/sleeve journal interface. The two primarybridges 29 are seen to incorporate hydraulic communication to cylinderports 4 and 5 via radial holes 28, a feature absent in the two secondarybridges 30.

Four styles of metering edge contours (henceforth termed "meteringedges") are employed on input-shaft 10, defining four types of orificesin the rotary valve: inlet and return orifices in the primary bridgeshenceforth termed primary inlet and primary return orificesrespectively, and inlet and return orifices in the secondary bridgeshenceforth termed secondary inlet and secondary return orificesrespectively.

FIG. 3 shows the upper half of FIG. 2 at a greater scale, and thereforeincorporates a single primary bridge 29 and a single secondary bridge30. Primary inlet orifices 31a, 31b are formed at the interface ofcoacting input-shaft grooves 18a and sleeve slots 19a. Primary returnorifices 32a, 32b are formed at the interface of coacting input-shaftgrooves 18b and sleeve slots 19a. Secondary inlet orifices 33a, 33b areformed at the interface of coacting input-shaft grooves 18a and sleeveslots 19b. Secondary return orifices 34a, 34b are formed at theinterface of coacting input-shaft grooves 18c and sleeve slots 19b.

The hydraulic "flow diagram" for this network of orifices is shown inFIG. 4. As can be seen from FIG. 2, the lower (hidden) half of therotary valve in FIG. 3 is axi-symmetric with respect to the upper halfand these halves function in parallel. The pump supply flow Q indicatedin FIG. 4 is therefore one half the total pump supply flow.

The manner of operation of the rotary valve will now be described inreference to the actual metering edge contours employed on the sides ofthe input-shaft grooves. These metering edge contours coact with theadjacent sleeve edges to generate the required restriction variationcharacteristic, as a function of valve operating angle θ.

Primary inlet orifices 31a, 31b are generated by primary inlet meteringedges 35, formed on one side of grooves 18a (refer to FIG. 5). Primaryreturn orifices 32a, 32b are generated by primary return metering edges36, formed on both sides of grooves 18b (refer to FIG. 6). Secondaryinlet orifices 33a, 33b are generated by secondary inlet metering edges37, formed on one side of grooves 18a opposite primary inlet meteringedge 35 (refer to FIG. 7). Secondary return orifices 34a, 34b aregenerated by secondary return metering edges 38, formed on both sides ofgrooves 18c (refer to FIG. 8).

FIGS. 5-8 depict the geometry of the four types of orifices at theneutral position of the rotary valve, designated θ=0 deg. In theforthcoming description, a clockwise rotation of input-shaft 10 withrespect to sleeve 11, numerically equal to a positive valve operatingangle θ, is considered to take place (refer to bold arrow in FIG. 3).Each pair of orifices of each type will therefore comprise one orificewhich is tending to close and one which is tending to open from thisneutral position. For example, referring to FIG. 6, primary returnorifice 32a tends to close, and eventually fully closes at θ=1.5 degwhen edge 39 of sleeve slot 19a reaches position 40. On the other hand,primary return orifice 32b tends to open further from this neutralposition and, for this same valve operating angle of θ=1.5°, edge 39reaches position 41. Note that, for simplicity in this description,relative angular rotation of input-shaft 10 and sleeve 11 isdiagrammatically shown as lateral motion of edge 39 with respect to afixed input-shaft metering edge.

FIG. 9 shows the angular boost characteristic of the rotary valve,expressed as differential pressure .increment.P on the Y axis plotted asa function of valve operating angle θ on the X axis. The aforementionedinput torque based boost characteristic, the basic "finger print" of arotary valve, is obtained by converting the X axis to input torque unitsby multiplying the abscissae by the torsional stiffness of torsion bar12. For example, for a torsion bar stiffness of 2 Nm/deg, a valveoperating angle of θ=4 deg corresponds to an input torque of 4×2=8 Nm.

The boost characteristic in FIG. 9 can be considered as comprising 3regions: an on-centre region 42 of low slope associated with on-centredriving, particularly high-speed on-centre driving under freewayconditions and, in this case, corresponding to valve operating angles upto about 1.5 deg; a cornering region 43 of medium slope associated withthe assistance pressures needed during vehicle cornering on windingcountry roads and, in this case, corresponding to valve operating angleof between about 1.5 deg and 4 deg; and a parking region 44 of highslope associated with the much larger assistance pressures required forstationary dry parking and, in this case, corresponding to valveoperating angles beyond about 4 deg. This boost characteristic is of theincreasing accepted style earlier referred to, namely an essentiallylinear cornering region 43 followed by a fast turn-around (as at point45) to parking region 44.

FIG. 10 shows the flow division between primary bridge 29 and secondarybridge 30 as a function of valve operating angle θ. FIG. 11 shows thissame relationship plotted as a function of differential pressure.increment.P.

Referring back to FIG. 4, in the neutral position of the rotary valvethe overall flow restriction provided by the orifices in secondarybridge 30, plus additional orifice 46 in series with secondary bridge 30(which will be described in detail later), is approximately three timesthe restriction offered by primary bridge 29. Flow Q therefore dividesin inverse proportion to this restriction resulting in approximately 75%of flow Q passing through primary bridge 29 ie. Qp/Q=0.75 in FIG. 10.Also because primary inlet orifices 31a and 31b are geometricallyequivalent in the neutral position, as are primary return orifices 32aand 32b, flow Qp evenly divides between flow Qpl in primary left-handlimb 47 and flow Qpr in primary right-hand limb 48, generating zerodifferential pressure .increment.P at cylinder 49.

Referring to FIG. 10, the diversion of 75% of flow to primary bridge 29is maintained more or less constant in on-centre region 42 of the boostcharacteristic, and is mainly a result of the relatively restrictivesecondary return orifices 34a, 34b. As seen in FIG. 8, secondary returnorifices 34a, 34b offer a substantially constant restriction area due tothroat 50, formed by the circumferential overlap of secondary returnmetering edge 38 and land 82 of sleeve bore 21. However because of therelatively unrestricted primary inlet orifices 31a, 31b and primaryreturn orifices 32a, 32b, and also the fact that all bridge limbs 47,48, 51 and 52 are open to flow in on-centre region 42, inlet pressure Pgenerated by the rotary valve is low under these conditions, henceaffording low energy losses in on-centre driving.

For increasing valve operating angle in on-centre region 42, primaryinlet orifice 31b and primary return orifice 32a progressively close,while primary inlet orifice 31a and primary return orifice 32bprogressively open, thereby maintaining the previous described conditionin which primary limb flows Qpl and Qpr are approximately equal, hencegenerating the low slope on-centre region 42 of the boostcharacteristic.

However, as the valve operating angle approaches 1.5 deg primary returnorifice 32a closes completely, as indicated by sleeve edge position 40in FIG. 6, diverting all primary flow Op down primary right-hand limb48. Simultaneously secondary return orifice 34a, which has also beenprogressively closing for increasing valve operating angle, now alsocloses completely, as indicated by slot edge position 53 in FIG. 8,diverting all secondary flow Qs down secondary right-hand limb 52. Now,as can be seen from FIG. 5, 6 and 7, opening orifices 31a, 32b and 33aoffer relatively little flow restriction at valve operating angles of1.5 deg or beyond hence, for the purposes of understanding the method ofoperation of the network of orifices, they can be ignored. Beyond valveoperating angles of 1.5 deg, the network shown in FIG. 4 can thereforebe considered to simplify to the arrangement shown in FIG. 12.Right-hand cylinder port 4 now effectively directly hydraulicallycommunicates to inlet port 2 and thus to the pump supply. Similarlyleft-hand cylinder port 5 now effectively directly hydraulicallycommunicates to return port 3 and thus to the pump reservoir.

In these circumstances all pump supply pressure (and hence valve inletpressure P) is applied to cylinder 49 (ie. P=.increment.P) and issubstantially determined by the restriction of the four remainingdominant orifices 31b, 33b, 34b and 46. The geometry of orifices 31b and33b are such that, as valve operating angle increases beyond 1.5 deg,primary inlet orifice 31b closes at a faster rate than does secondaryinlet orifice 33b, thereby diverting primary flow Qp to the secondarybridge and hence correspondingly increasing Qs. Thus, in cornering zone43, as is evident in FIGS. 10 and 11, the flow ratio Qp/Q progressivelyreduces from about 0.75 at a valve operating angle of 1.5 deg andeventually reaches 0 (zero) at a valve operating angle of 4 deg , whereprimary inlet orifice 31b fully closes. The geometry of orifices 31b and33b therefore predominantly determine the shape of the boostcharacteristic in cornering zone 43, in this case a linear boostcharacteristic. In cornering zone 43, as differential pressure P buildsup and is directly applied across orifice 31b, flow is simultaneouslyprogressively diverted away from orifice 31b according to the mechanismdescribed above. For example, looking at FIG. 11 it is seen that Qp hasdropped to about one-half its on-centre value when differential pressureP reaches 1 MPa. This action is arranged such that orifice 31b neverproduces any substantial cavitation noise since, as is well known in theart, valve cavitation noise generated in a given orifice reduces withrate flow through the orifice for a given fixed pressure drop.

Now, in the absence of orifices 34b or 46, the corresponding increase insecondary flow Qs through orifice 33b would certainly cause this orificeto produce cavitation noise. This increase in noise would not only becaused by the increase in secondary flow Qs but also the increasingrestriction of orifice 33b for increasing valve operating angle. Howeverincrease in secondary flow Qs, for example by the factor of fourexampled by this embodiment (refer to FIGS. 10 and 11), dramaticallyincreases the back pressures Pb and Pr generated by orifices 34b and 46respectively. In the embodiment shown Pb and Pr are arranged to reach 1MPa and 200 kPa respectively when secondary flow Qs reaches its maximumvalue. This rise in back pressures Pb and Pr is demonstrated graphicallyin FIGS. 13 and 14.

Once orifice 31b has fully closed off at the end of cornering zone 43,all pump flow now passes through orifices 33b, 34b and 46 in series.Hence back pressures Pb and Pr are held constant in parking zone 44,which extends from valve operating angles 4 deg to 4.5 deg and in whichdifferential pressure correspondingly rises from 2 MPa to 8 MPa. Sharpturnaround 45 between cornering zone 43 and parking zone 44 is aided bythe total diversion of flow to secondary bridge 30 and hence to orifice33b, plus the relatively steep "close-off angle" of metering edge 37 asat region 60 (refer to FIG. 7).

In this manner the back pressure developed by orifices 34b and 46 inseries (ie. Pr+Pb) "tracks" (or follows) the increase in pressuredeveloped across potentially cavitating orifice 33b during corneringregion 43 ie. up to a maximum differential pressure P of 2 MPa. At thistime 1.2 MPa of this 2 MPa is actually attributable to back pressureorifices 34b and 46 (refer to FIGS. 13 and 14). This back pressure isthen held constant at 1.2 MPa for the remaining parking zone 44 duringwhich time differential pressure rises to 8 MPa. The 1.2 MPa backpressure generated at the exit orifice 33b is sufficient to suppresssubstantially all cavitation noise from this orifice, even up to themaximum differential pressure of 8 MPa used for parking.

Referring to FIG. 8, it is seen that orifice 34b is generated bymetering edge 38 which circumferentially overlaps the adjacent land 82of sleeve bore 21 for all valve operating angles from the neutralposition up to the maximum valve operating angle of 4.5 degcorresponding to slot edge position 61. Radially disposed "throat" (orpoint of minimum cross-sectional area to flow) 50 serves to ensure thatorifice 34b provides a substantially constant restriction area to oilflow through this range of valve operating angles and beyond, 7 deg inthis case matching the fail-safe angle of the rotary valve wheremechanical stops at the interface of input-shaft 10 and pinion 7 preventany further relative rotation between input-shaft 10 and sleeve 11.

The shape of metering edge 38, including region 62 (of locally reducedmetering edge depth which forms throat 50 in combination with land 82 ofsleeve bore 21), also aids in smoothing the normally turbent oil flow asit passes sleeve slot edge 63. For less demanding applications wherethis turbulent flow problem does not necessarily propagate as valvenoise, metering edge 38 can be made with a simpler flat bottom form 64(ie. a substantially constant metering edge depth), still providing anapproximately constant restriction area for orifice 34b.

If residual cavitation noise is generated at sleeve slot edge 63, backpressure can be applied to orifice 34b via the presence of downstreamfixed orifice 46. In this first embodiment of the first aspect of thepresent invention, secondary return holes 25 in input-shaft 10 (see FIG.3) are of reduced diameter compared to primary return holes 24, andproduce a back pressure Pr of 200 kPa at maximum secondary flow Qs(refer to FIGS. 13 and 14). Turbulence generation in these holes can bereduced (if necessary) if a conical (FIG. 15a), recessed conical (FIG.15b), axi-symmetric convex tapered (FIG. 15c) or recessed axi-symmetricconvex tapered (FIG. 15d) entry to holes 25 is employed. Such entryprofiles, and numerous others, can be readily machined via a "steppeddrill" arrangement to smoothen the inlet flow to holes 25.

Orifices 34b and 46 in series provide a staged pressure reductiondownstream of orifice 33b, enabling large back pressures to be appliedto this secondary inlet orifice without generation of any substantialcavitation noise. For example, at maximum parking differential pressureof 8 MPa, the pressure drop generated by orifice 33b is 6.8 MPa, thepressure drop generated by orifice 34b is 1 MPa and the pressure dropgenerated by orifice 46 is 200 kPa. An important feature of the presentinvention is that this by-pass mode valve arrangement enables backpressure Pb+Pr to be applied directly to cylinder 49 for all valveoperating angles beyond 1.5 deg (ie. throughout cornering zone 43 andparking zone 44) corresponding to the region of close-off of primaryreturn orifice 32a. This is particularly beneficial in parking zone 44corresponding to valve operating angles beyond 4 deg where, primaryinlet orifice 31b having now closed off, back pressure Pb+Pr reaches itsmaximum value of 1.2 MPa. This relatively high magnitude of backpressure is used usefully to produce a force on the piston in cylinder49, rather than wastefully dissipating energy as heat.

Input-shaft metering edges 35, 36, 37 and 38 can be readily manufacturedusing coining or roll-imprinting processes well known in the art. Suchrelatively steep (ie high slope with respect to adjacent land 81 ofinput-shaft 10) metering edges enable good control of the boostcharacteristic, hence steering effort levels, and according to thepresent invention has the potential to reduce noise levels in the rotaryvalve to less than 55 dBA.

FIGS. 16, 17, 18, 19a-d and 20 show a second embodiment of the firstaspect of the present invention in which fixed orifices 46, rather thanbeing generated by reduced diameter radial holes 25, are generated by arestriction to axial oil flow within bore 26 of input-shaft 10.Secondary return holes 70 in this second embodiment are not intended tobe restrictive, but serve to communicate hydraulic oil from secondarybridges 30 to bore 26 of input-shaft 10 whence oil flow turnsperpendicularly to flow axially along bore 26 (refer to FIG. 16).However prior to reaching return port 3, the flow must pass throughannular restriction 71 formed by the interaction of diametricallyenlarged portion 72 of torsion bar 73 and reamed portion 74 ofinput-shaft bore 26. This annular restriction 71 constitutes the twofixed orifices 46, previously described in reference to the firstembodiment of the present invention. According to this secondembodiment, hydraulic oil from primary bridge 29 is communicated moredirectly to return port 3 (refer to FIG. 17) so that it is not requiredto flow through bore 26 of input-shaft 10, and hence is not subject toannular restriction 71. This is achieved by axially extending grooves18b of input-shaft 10 associated with primary return orifices 32a,b toform axially disposed channels 75. Channels 75, two of which arerequired according to this second embodiment, can be readilymanufactured by processes well known in the art such as milling orplunge grinding. Note that at least one of these two channels 75 must beextended axially in the opposite direction to the main flow direction inorder to bleed leakage oil from the cavity on the input side (right sidein FIG. 17) of sleeve 11.

FIGS. 19a-d show various possible embodiments for diametrically enlargedportion 72 of torsion bar 73.

FIG. 19a shows in more detail diametrically enlarged portion 72 oftorsion bar 73, forming annular restriction 71 via its interaction withreamed (or otherwise accurately diametrically sized) portion 74 of bore26 of input-shaft 10. The general direction of return oil flow fromsecondary bridges 30 is shown by large arrows and, as seen, annularrestriction 71 constitutes orifices 46 and generates the back pressurePb of 200 kPa previously referred to. Because of the largecircumferential length of annular restriction 71 (typically 20-30 mm),flow noises in certain applications will be less than flow noisesgenerated by reduced diameter radial holes 25 employed in the firstembodiment of the present invention.

As shown in FIG. 19b, restriction 71 may also be staged, that isdiametrically enlarged portion 72 comprises a series of axiallyseparated, circumferentially disposed lands. This staging means that theback pressure generated by restriction 71 is generated in a series ofdiscrete stages, in the case of FIG. 19b in three stages. The majorbenefit of this arrangement, compared to the simple cylindrical form ofenlarged portion 72 of torsion bar 73 shown in FIG. 19a, is that theback pressure generated by staged restriction 71 will be less variableas a function of oil viscosity and hence oil temperature. This isbecause most of the back pressure generation occurs around the sharpedges associated with the circumferentially disposed lands.

In applications where torsion bar 73 is stressed to its maximumendurance limit, diametrically enlarged portion 72 may be plasticmoulded around the otherwise conventional torsion bar as a separate,subsequent operation as shown in FIG. 19c. In this embodiment theoutside diameter of diametrically enlarged portion 72 is fluted as at 79(refer to FIG. 20) and interference-fitted into reamed portion 74 ofbore 26, thereby ensuring the radial depth accuracy of annularrestriction 71. In this embodiment, plastic moulded diametricallyenlarged portion 72 can also be arranged to generate a stagedrestriction similar to that shown in FIG. 19b.

FIG. 19d shows another embodiment where diametrically enlarged portion72 is extended axially to overlap secondary return hole 70. The flow isalso smoothened by employing conical portion 76 to redirect the flowfrom the radial direction in hole 70 to the axial direction in bore 26.The use of a plastic material surrounding torsion bar 73 has also beenfound to assist in the dampening of turbulence noise generated when theradial oil flow through hole 70 impinges on torsion bar 73.

To this end, it is also possible to employ a plastic moulding 77 aroundtorsion bar 12 according to the first embodiment of the first aspect ofthe present invention. In this latter case diametrically enlargedportion 72, provides an additional back pressure generating capabilityand hence augments reduced diameter radial drill holes 25. Thisarrangement is shown in FIG. 21 where optional conical portion 78 servesthe same function as conical portion 76 in the second embodiment of thefirst aspect of the present invention.

FIGS. 22, 23 and 24 show a third embodiment of the first aspect of thepresent invention in which fixed orifices 46 are generated by arestriction to axial oil flow outside bore 26 of input-shaft 10, indeedthe secondary return path does not pass through bore 26. Axiallyextending primary return channels 90 and secondary return channels 91are preferably broached in sleeve bore 21 and are respectivelycircumferentially aligned with (and hence hydraulically communicatewith) primary return grooves 18b and secondary return grooves 18crespectively, thereby communicating hydraulic fluid to return port 3without the need for hydraulic communication to bore 26 of input-shaft10. The radial depth of secondary return channels 91 is small comparedto their width, hence creating a high aspect ratio restriction 46 in thesecondary return path via its interaction with adjacent input-shaftoutside diameter 20. The radial depth of primary return channels 90 isrelatively large compared to the depth of secondary return channels 91,the former therefore generating minimum restriction in the primaryreturn path.

In a fourth embodiment of the first aspect of the present invention (notshown as a separate figure), secondary return grooves 18c can be axiallyextended as shallow, high aspect ratio channels formed via theirinteraction with adjacent sleeve bore 21. Such channel-like extensionsof secondary return grooves 18c can be readily achieved via milling orgrinding operations well known in the art, and will appear similar tochannel 75 shown in FIG. 17 in reference to another earlier embodimentexcept that the radial depth would need to be less in order to generatethe necessary restriction 46 in the secondary return path. Primaryreturn grooves 18b can be similarly axially extended as a radiallydeeper channel to facilitate a relatively unrestricted primary returnpath to return port 3.

Both third and fourth embodiments of the first aspect of the presentinvention offer the advantage that, if both primary and secondary returnpaths are facilitated with channels, no drill holes are required in theinput-shaft. This feature potentially simplifies and reduces cost ofmanufacture of this component.

FIGS. 29, 30, 31 and 32 show a fifth embodiment of the first aspect ofthe present invention in which previously referred to fixed orifices 46are generated by annular restriction 100 formed at the interface ofinput-shaft 10 and sleeve 11. Secondary return grooves 18c are axiallyextended to communicate with annular cavity 101 formed by theinteraction of reduced diameter portion 102 of the outer periphery ofinput-shaft 10 and bore 82 of sleeve 11. The annular cavity acts as amanifold to gather secondary return flow, which is then communicated viaannular restriction 100 to return port 3. Annular restriction 100 ispreferably generated by a predetermined small radial clearance existingbetween reduced diameter portion 102 of input-shaft 10 and the insidediameter 103 of radially inwardly extending annular pressed metal cup104, press fitted inside sleeve skirt 105 to seal against the axialextremity of sleeve bore 82. As is seen in FIGS. 31 and 32, the primaryreturn path passes through bore 26 of input-shaft 10 in a similar mannerto that described in reference to the first embodiment, thereby notbeing subject to annular restriction 100.

According to the second aspect of the present invention, an annularrestriction to secondary return oil flow replaces a major part of thefunction of secondary return orifices 34a,b disclosed in the firstaspect of the present invention. The secondary return orifices in thisembodiment need not provide the substantially constant restriction area(provided by throat 50) referred to in embodiments according to thefirst aspect of the present invention, this function now being entirelyprovided by the annular restriction. According to both embodiments ofthe second aspect of the present invention, secondary return orifices34a,b are 1.5 deg close-off orifices whose function is controlled bymuch simpler metering edge 83 (refer to FIG. 25). Primary inlet orifices31a,b, primary return orifices 32a,b and secondary inlet orifices 33a,bremain unchanged from the first aspect of the present invention (referback to FIGS. 5, 6 and 7). Primary return orifices 32a,b and secondaryreturn orifices 34a,b are therefore identical in geometry in theseembodiments of the second aspect of the present invention, quitedistinct from embodiments described in reference to the first aspect ofthe present invention.

FIGS. 26, 27 and 28 show more general views of the first embodiment ofthe second aspect of the present invention. It is seen that, as in thecase of the second embodiment of the first aspect of the presentinvention, back pressure in the secondary return path is generated byannular restriction 71 formed by diametrically enlarged portion 72 oftorsion bar 73 and its coaction with accurately diametrically sizedregion 74 of input-shaft bore 26. FIGS. 26 and 27 are in fact identicalto FIGS. 16 and 17 respectively. FIG. 28 is similar to FIG. 18, exceptfor the differing geometry of secondary return orifices 34a,b. All otheraspects of this embodiment of the second aspect of the present inventioncan be considered as being according to FIGS. 19 and 20, alreadydescribed in reference to the second embodiment of the first aspect ofthe present invention.

FIGS. 33, 34, 35 and 36 show a second embodiment of the second aspect ofthe present invention in which back pressure in the secondary returnpath is generated by annular restriction 100 formed at the interface ofinput-shaft 10 and sleeve 11. Secondary return grooves 18c are axiallyextended to communicate with annular cavity 101 formed by theinteraction of reduced diameter portion 102 of the outer periphery ofinput-shaft 10 and bore 82 of sleeve 11. The annular cavity acts as amanifold to gather secondary return flow, which is then communicated viaannular restriction 100 to return port 3. Annular restriction 100 ispreferably generated by a predetermined small radial clearance existingbetween reduced diameter portion 102 of input-shaft 10 and the insidediameter 103 of radially inwardly extending annular pressed metal cup104, press fitted inside sleeve skirt 105 to seal against the axialextremity of sleeve bore 82. As seen in FIGS. 35 and 36, the primaryreturn path passes through bore 26 of input-shaft 10 in a similar mannerto that described in reference to the first embodiment of the firstaspect of the present invention, thereby not being subject to annularrestriction 100.

According to these first and second embodiments, in the absence of anysubstantial back pressure generation by secondary return orifices 34a,bwhen secondary inlet orifices 33a,b are closing, all back pressure tosuppress cavitation noises in these latter orifices must necessarily beprovided by annular restrictions 71 and 100 respectively. These annularrestrictions must therefore now supply a back pressure up to 1.2 MPa ormore, as opposed to the 200 kPa (say) required to be supplied by orifice46 according to the respective second and fifth embodiments of the firstaspect of the present invention.

In these circumstances noise generation in annular restrictions 71 and100 is minimised by their high aspect ratio, such terminology in thepresent specification meaning the ratio of the general proportions ofthe cross-section of the restriction, this ratio always expressednumerically as unity or greater. In the case of the first embodiment,this aspect ratio is more specifically the circumferential length ofannular restriction 71 (ie. approximately πD where D is the diameter ofdiametrically enlarged portion 72) divided by the radial depth of therestriction (ie. the radial clearance between the outside diameter ofdiametrically enlarged portion 72 and accurately diametrically sizedregion 74). In the case of the second embodiment, this aspect ratio isthe circumferential length of annular restriction 100 (ie. approximatelyπD where D is the diameter of reduced diameter portion 102) divided bythe radial depth of the restriction (ie. the radial clearance betweenreduced diameter portion 102 and inside diameter 103). In order thatrestrictions 71 and 100 in these embodiments do not generate cavitationnoise, an aspect ratio of greater than 50 has been found to be required.However, for certain applications where lesser levels of back pressuremay be required to be developed by these annular restrictions, aspectratios as low as 10 may be still practical.

The staging of restriction 71 shown in FIG. 19b will provide aparticular benefit in the first embodiment because it serves todramatically reduce the viscosity sensitivity of this restriction, whichis now of relatively small cross-sectional area in order to generate thelarger amount of back pressure. The staging of annular restriction 71also assists in minimising noise generation in this orifice.

It will be recognised by persons skilled in the art that numerousvariations and modifications may be made to the invention withoutdeparting from the spirit or scope of the invention.

I claim:
 1. A rotary valve for a hydraulic power steering gearcomprising a value housing having an inlet port to receive hydraulicfluid from a pump, a return port to return hydraulic fluid to the pump,and cylinder ports to communicate hydraulic fluid to left and right-handcylinder chambers of the power steering gear, the valve also comprisingan input shaft having in its outer periphery a plurality of axiallyextending grooves separated by lands, a sleeve journalled on saidinput-shaft and rotationally secured to a driven member, said sleevehaving in its bore an array of axially extending slots circumferentiallyaligned with the lands on the input-shaft, the interfaces between theco-acting input-shaft grooves and sleeve slots defining axiallyextending orifices controlling fluid flow within the valve, the orificesopening and closing when relative rotation occurs between theinput-shaft and sleeve from a neutral position, a torsion bar residingin a bore of the input-shaft compliantly connecting the input-shaft anddriven member, and arranged to urge the sleeve and input shaft to theneutral position, the orifices being ported as a network such that theyform at least one primary and at least one secondary hydraulicWheatstone bridge arranged in parallel, each said bridge comprising twolimbs hydraulically communicating the inlet and return ports, each saidlimb containing an inlet orifice hydraulically communicating to theinlet port and a return orifice hydraulically communicating to thereturn port, the magnitude of the hydraulic flow through each bridgevarying in accordance with the restriction offered by the respectiveinlet and return orifices in that bridge, the limbs of the primarybridge incorporating means providing hydraulic communication to one ofthe cylinder ports at a point of interconnection of the respective inletand return orifices in that limb, the limbs of the secondary bridge notincorporating means providing hydraulic communication to the cylinderports, wherein the hydraulic flow from said primary bridge ishydraulically communicated to the return port via a primary return pathand the hydraulic flow from the secondary bridge is hydraulicallycommunicated to the return port via a secondary return path, an annularrestriction existing in the secondary return path, said annularrestriction having a cross-section to flow which has a high aspectratio.
 2. A rotary valve as claimed in claim 1, wherein said annularrestriction has a cross-section to flow which has an aspect ratiogreater than
 10. 3. A rotary valve as claimed in claim 1, wherein one ofthe primary return path and the secondary return path passes through thebore of the input-shaft and the other of the primary return path andsecondary return path does not pass through the bore of the input shaft.4. A rotary valve as claimed in claim 3, wherein the secondary returnpath passes through the bore of the input-shaft and the annularrestriction is formed within this bore.
 5. A rotary valve as claimed inclaim 4, wherein the annular restriction formed in the bore of theinput-shaft is generated by virtue of a small radial clearance existingbetween a diametrically enlarged portion of the torsion bar and theinput-shaft bore, thereby restricting axial flow in the bore towardssaid return port.
 6. A rotary valve as claimed in claim 5, wherein saiddiametrically enlarged portion of the torsion bar is formed as a plasticmoulded annular bush around said torsion bar.
 7. A rotary valve asclaimed in claim 6, wherein said plastic moulded annular bush is made ofan engineering plastic chemically resistant to hydraulic oil.
 8. Arotary valve as claimed in claim 7, wherein the engineering plastic isDelrin®.
 9. A rotary valve as claimed in claim 7, wherein theengineering plastic is Lurathane®.
 10. A rotary valve as claimed inclaim 6, wherein said plastic moulded annular bush extends axially tooverlap secondary return radial holes in said input-shaft.
 11. A rotaryvalve as claimed in claim 4, wherein hydraulic flow from the primarybridge is directly communicated to the return port via channels formedas axial extensions of the input-shaft grooves associated with theprimary return orifices, the hydraulic flow thereby not being subjectedto the annular restriction.
 12. A rotary valve as claimed in claim 3,wherein the secondary return path does not pass through the bore of theinput-shaft and the annular restriction is formed at theinput-shaft/sleeve interface.
 13. A rotary valve as claimed in claim 12,wherein the secondary return grooves are axially extended in a least onedirection to communicate with an annular cavity formed by theinteraction of a reduced-diameter portion of the input-shaft outerperiphery and the sleeve bore, said annular cavity acting as a manifoldto gather secondary return oil flow, which is then communicated via anannular restriction to the return port.
 14. A rotary valve as claimed inclaim 13, wherein said annular restriction is generated by apredetermined small radial clearance existing between said reduceddiameter portion of the input-shaft and the inside diameter of aradially inwardly extending portion of the sleeve bore.
 15. A rotaryvalve as claimed in claim 14, wherein said radially inwardly extendingportion of the sleeve bore is formed as an accurately internally andexternally sized annular pressed metal cup which is press-fitted insidethe sleeve skirt to seal against the axial extremity of the sleeve bore.16. A rotary valve as claimed in claim 12, wherein said primary returnpath passes through the bore of the input-shaft thereby bypassing theannular restriction en-route to the return port.
 17. A rotary valve asclaimed in claim 1, wherein hydraulic flow from the secondary bridgepasses axially through said annular restriction and hence applies a backpressure to all secondary orifices upstream of said restriction, therebyproviding sufficient back pressure to suppress the generation ofcavitation noise to said secondary orifices for all valve operatingangles.
 18. A rotary valve as claimed in claim 1, wherein said rotaryvalve has eight input-shaft grooves.
 19. A rotary valve as claimed inclaim 1, wherein said rotary valve has eight sleeve slots.
 20. A rotaryvalve for a hydraulic power steering gear comprising a value housinghaving inlet port means for receiving hydraulic fluid from a pump,return port means for returning hydraulic fluid to the pump, andcylinder port means for communicating hydraulic fluid to left andright-hand cylinder chambers of the power steering gear, the valve alsocomprising an input-shaft and a sleeve journalled on the input-shaft androtationally secured to a driven member wherein the input-shaft andsleeve define axially extending orifice means for controlling fluid flowwithin the valve, the orifice means opening and closing when relativerotation occurs between the input-shaft and sleeve from a neutralposition, a torsion bar residing in a bore of the input-shaftcompliantly connecting the input-shaft and driven member, and arrangedto urge the sleeve and input shaft to the neutral position, the orificemeans being ported as a network such that they form at least one primaryand at least one secondary hydraulic Wheatstone bridge arranged inparallel, each said bridge comprising two limbs hydraulicallycommunicating the inlet and return port means, each said limb containinginlet orifice means for hydraulically communicating to the inlet portmeans and return orifice means for hydraulically communicating to thereturn port means, the magnitude of the hydraulic flow through eachbridge varying in accordance with the restriction offered by therespective inlet and return orifice means in that bridge, the limbs ofthe primary bridge incorporating means for providing hydrauliccommunication to one of the cylinder port means at a point ofinterconnection of the respective inlet and return orifice means in thatlimb, the limbs of the secondary bridge not incorporating means forproviding hydraulic communication to the cylinder port means, whereinthe hydraulic flow from said primary bridge is hydraulicallycommunicated to the return port means via a primary return path and thehydraulic flow from the secondary bridge is hydraulically communicatedto the return port means via a secondary return path, an annularrestriction existing in the secondary return path, said annularrestriction having a cross-section to flow which has a high aspectratio.